Different types of rotary valves for internal combustion engines have been known since the early days of conception of such engines as an alternative to poppet type valves to selectively and periodically enable and prevent flow of intake fluids and exhaust gases through a transfer port leading to a combustion chamber of the engine.
The potential advantages which use of rotary valves can provide as compared to conventional poppet type valves are well documented and include improved smoothness of operation, rapid and precise opening and closing of transfer ports, reduction or inhibition of intake-exhaust overlap and the larger port opening sizes that can be provided to achieve higher volumetric ratios of fluid transfer into/from the combustion chamber. The major drawback in their practical implementation is the hitherto unsatisfactorily resolved sealing problems.
In the following, rotary valves will be described in the context of their use with reciprocating type internal combustion engines. However, neither some of the known nor the rotary valve embodiments in accordance with the present invention are limited to such applications. They may find equal use with other type of engines, such as rotary piston engines (Wankel motor), in which combustion energy is transformed into mechanical power in an intermittent operating cycle of the engine. The operating cycle, as used herein, encompasses the induction of a working fluid (air, fuel) into, compression and subsequent ignition of the working fluid within, expansion of resultant combustion gases within and exhaustion of combustion gases from a combustion chamber of the engine. These are also referred to below as the operating phases or strokes (in the case of reciprocating type engines) of the engine.
One such rotary valve type broadly consists of a cylindrical shaped rotor body which is coaxially supported for rotation within a valve bore formed in the cylinder head of the engine. Gas exchange ports formed on the peripheral surface of the rotor body periodically align with an associated transfer port opening in the bore and which leads to a combustion chamber of the engine. Examples of such valves include valve rotors with radial gas flow only, in which one or more gas exchange ducts have diametrically opposing ports in an otherwise continuos peripheral surface (see eg U.S. Pat. No. 4,019,499), and valve rotors with "axial--radial gas flow" having two gas exchange ducts commencing on opposite axial side faces of the valve rotor and extending therethrough so as to respectively terminate in an intake and an exhaust port in the peripheral surface of the rotor body (see eg U.S. Pat. No. 5,052,349). Rotary valves can be equally employed for two and four stroke engines, specific layout of the gas transfer ducts and ports on the rotor being also dependant on the operating speeds of the rotor with respect to the crankshaft.
It is necessary to emphasise at the outset that the present invention is concerned with those types of rotary valve constructions in which the cylindrical rotor body is coaxially supported for rotation within the valve bore so as to maintain a relatively "small" radial clearance gap between the bore surface and the peripheral rotor body surface; "small" in this context is a 0.2 to 0.4 mm radial clearance gap which will ensure rotation of the rotor without risk of seizure within the bore and allows for manufacturing tolerances that are achievable without excessive costs.
Such rotary valves require a "seal system" arranged to define a frame around the transfer port which bridges and closes the radial gap so as to minimise gas leakage from the transfer port while the latter is to be maintained closed by the rotary valve, in particular during the ignition phase. This class of rotary valves is exemplified by the one disclosed in U.S. Pat. No. 4,852,532 ("the first Bishop patent"). The contents of this first Bishop patent is included herein by way of short hand cross reference, in particular in so far as it contains a succinct evaluation of some relevant prior rotary valve types and their relative drawbacks, see in particular column 3, line 56 to column 4, line 62.
In the rotary valve construction proposed in the first Bishop patent, a sleeve-like rotor has separate intake and exhaust ducts beginning in opposite axial side faces of the rotor body. The ducts respectively terminate in an inlet and exhaust port angularly spaced apart on the peripheral surface of the rotor body at the same axial location. The ports are dimensioned such that upon rotation of the valve rotor within the valve bore formed in the cylinder head of the engine, the inlet and exhaust ports periodically align with and pass over a single transfer port communicating the bore with the combustion chamber defined within the cylinder. The location and circumferential extension of the peripheral ports and herein between extending rotor surface zone, termed in Bishop "sealing zone", are chosen and dimensioned such that given a properly timed rotor revolution speed with respect to the operating cycle of the engine, gas passage to/from the transfer port is enabled through the intake and exhaust ports of the rotor body and prevented while the "sealing surface" covers the transfer port. In other words, specific or "discrete" zones on the periphery of the rotor body (including those surface zones which contain the ports) are "associable" with a respective one of the phases of the operating cycle, eg the same "discrete surface zone" of the rotor body always passes over and covers the transfer port during the exhaust, intake and combined compression and expansion phases.
In the first Bishop patent is disclosed a system of so-called "floating seals" which consists of two longitudinal sealing elements, rectangular in cross-section, which are received in grooves extending parallel with respect to the axis of rotation of the rotor and formed on either circumferential side of the transfer port within the bore surface. This axially extending sealing elements are loaded against the peripheral surface of the rotor body portion which includes the gas exchange ports, thereby to bridge the radial gap and prevent gas flow from the transfer port past the seals in circumferential direction of the rotor body. Such seals will hereinafter also be referred to as "seal elements against circumferential flow".
The seal elements against circumferential flow are abutted at either longitudinal end at a sealing ring respectively received within an annular groove formed on either axial side of the transfer port within the bore surface. The radially inward facing inner peripheral surface of the sealing rings sealingly rubs against the peripheral surface of the rotor body thereby to prevent gas flow from the transfer port past the sealing rings in axial direction of the rotor body. Such sealing rings will herein after also be referred to as "seal elements against axial flow".
The main function of the "floating seal frame" disclosed in the first Bishop patent is to prevent leakage of high pressure combustion gases primarily created during and subsequent the ignition phase of the operating cycle of the engine into the radial gap volume outside the framed transfer port, and thereby into the gas exchange ports of the rotor body and the axially adjoining rotor zones at which the rotor is supported in roller bearings. The effectiveness of the sealing system depends on the ability of the sealing elements against circumferential and axial flow to maintain a "closed" frame in particular during the critical compression, ignition and expansion phases. With the "seal frame" of the first Bishop patent this is not possible.
Due to manufacturing tolerances, assembly requirements and the nature of the sealing system employed, the width and depth of the annular grooves and of the longitudinal grooves will always need to be greater than the width and depth, respectively, of the respective sealing elements intended to be received therein. Thus, in case of the seal elements against circumferential flow, these will be received between the radially extending side surfaces of the longitudinal grooves with predetermined play in circumferential direction. This play is in itself not critical, since high pressure compression and combustion gases will tend to load the longitudinal seal elements in circumferentially opposite directions into sealing abutment against the groove side surfaces farthest from the transfer port, thereby creating an effective sealing band along the longitudinal groove.
However, in case of the seal elements against axial flow, the high pressure gases will bias the sealing rings in axially opposite directions away from the transfer port and against the radially extending side surfaces of the annular grooves farthest from the transfer port. This opens up the already existing gap between the longitudinal ends of the longitudinal seal elements and the hereto adjacent sealing rings, creating a "leakage path" for gases and fluids. Bearing in mind that the sealing rings maintain a radial clearance gap to the bottom of the annular grooves, a relevantly large leakage path cross section is created at the four seal element intersection points of the sealing frame. This leakage path cross section for each intersection is given by the product of the axial clearance gap between the axial end of the longitudinal sealing element and the respectively adjacent sealing ring (which in most cases would be equal to the difference between annular groove width and sealing ring width), and the radial extension (depth) of the sealing ring plus the product of the radial clearance gap between the bottom of the annular groove and the inner circumferential surface of the sealing ring and the width of the annular groove.
It can be demonstrated that the gas leakage rate from the combustion chamber past conventional piston sealing rings to the crankcase housing is directly proportional to the leakage path area given by the product of ring gap and radial clearance of the piston crown to the cylinder bore diameter. It can be further shown that the total leakage path area described above, on the basis of reasonable assumptions as to the clearance and tolerance values for above elements, to be in the order of twenty times the leakage path area of a conventional piston ring assembly of the piston reciprocating in the cylinder for which the rotary valve is to serve as closure means during the mentioned operating cycle phases of the engine. Thus, the overall gas leakage rate from the combustion chamber stemming from sealing system inadequacies of the rotary valve is potentially quite larger than is the case with conventional poppet type valves, where no such additional leakage is present.
This leakage in turn has a number of adverse effects on engine performance values. In case of an engine with carburettor type fuel delivery system, noticeable amounts of unburnt fuel can leak during the end phases of the compression stroke past the leakage zones in circumferential direction along the annular grooves into the radial gap outside the framed transfer port and into the exhaust port, thereby producing unwanted hydrocarbon emissions.
Further, the combustion flame proper can expand into the "crevices" underneath the seal elements (radial clearance gap which is always maintained between groove bottom and seal element) and between the seal elements when these are forced away at the intersection points. This not only can lead to unwanted combustion deposits on the sealing surfaces, which adversely affect the seal system in time, but also to flushing of only partly combusted charge mixture into the exhaust manifold system.
The above mentioned problems have been recognised and sought to tie addressed in PCT patent publication WO 94/11618 by Bishop (the second Bishop patent). There, a total of four sealing elements against axial flow and two sealing elements against circumferential flow are disposed about surround the transfer port. The sealing elements against circumferential flow are again received in corresponding longitudinal grooves formed in the bore surface on either circumferential side of the transfer port. In contrast to the first Bishop patent, the sealing elements against axial flow (sealing rings) are received in annular grooves formed in the circumferential surface of the rotor body on either axial side of the exchange ports. The sealing rings are preloaded with their radially outward facing circumferential surface to rub against the bore surface.
In one embodiment of the second Bishop patent, one annular groove is provided at either axial end of the rotor and two sealing rings are received with small axial play between them in each groove. Further, the sealing rings located closest to the transfer port have an arc segment with reduced depth. The length of this segment is equivalent to the distance in circumferential direction between the longitudinal sealing elements. That is, the sealing ring has an arc portion having an outer diameter which is smaller than the outer diameter in the non-recessed arc portion. The reduction in depth is by an amount equal to or greater than the radial clearance gap between the bore surface and the cylindrical main portion of the rotor, so that this arc portion does not rub against the bore surface. Rectangular indentations are disposed at each circumferential end of the reduced depth arc segment to accommodate the axially opposite ends of the longitudinal sealing elements; thus, these elements no longer abut with their terminal end faces on the radially extending side faces of the inner most rings, as is the case with the seals of the first Bishop patent, but rather the axially extending side faces that face away from the transfer port will circumferentially engage against the radially extending surfaces of the indentations to effect sealing.
This sealing system relies on the compression and in particular the ignition pressure to seal off gas leakage from the transfer port past the seals in axial and circumferential direction of the rotor by "subdividing" sealing functions. Instead of entirely confining high pressure gases into the rectangular transfer port zone framed by the "window of floating seals" of the first Bishop patent, these gases are now allowed to expand into the annular pressurising cavities defined at either axial end of the longitudinal seal elements and formed between the facing axial side surfaces of the two sealing rings, the outer circumferential surfaces of the sealing rings received in the groove, the side and bottom surfaces of the groove within which each ring pair is received with radial clearance to the groove bottom, and the surface of the bore against which the inner circumferential surfaces of the sealing rings rub. This expansion takes place through the reduced depth arc segment of the axially inner sealing rings. The high pressure biases the rings in axially opposite directions into sealing contact with the side surfaces of the annular groove as well as in radially outward direction thereby augmenting contact pressure of the outer circumferential surface of each ring against the bore surface.
The longitudinal seal elements are, as is the case in the first Bishop patent, biased in circumferentially opposite directions by the high pressure compression and combustion gases into sealing abutment against the groove side surfaces farthest from the transfer port, thereby creating an effective sealing band along the longitudinal groove with the exception of the intersection points between inner sealing rings and longitudinal sealing elements. This sealing mechanism is, therefore, intended to minimise circumferentially directed gas leakage along the rotor body surface toward the gas exchange ports.
Admittedly, the above described sealing system in accordance with the second Bishop patent suppresses gas leakage from the transfer port in axial outward direction past the axially outer, piston ring-type, sealing rings to that present at the ring gap. However, due to manufacturing tolerances and the inability to precisely juxtapose the longitudinal sealing elements (which are housed in the stationary bore) with respect to the axially inner sealing rings (which are housed in the rotating rotor), gas leakage gaps between the side surfaces of the longitudinal sealing elements received within the rectangular indentations of the inner sealing rings will still be present.
Of greater concern is that highly corrosive combustion gases and, as currently believed, combustion flames are permitted to enter the crevices formed by the annular pressurising cavities. The problems of potential deposits within the crevices and on surfaces intended to perform sealing functions is therefore still unaddressed.